Centrifugal impeller



Oct. 11, 1949.. c. CONCORDIA ET AL CENTRIFUGAL IMPELLER 2 Sheets-Sheet 1Filed Dec. 20, 1945 Inventor-s: Gh ar-les Concordia,

Millard F. Dowell. 10 5 mm Their Attorngy.

Oct. 11, 1949.

Filed Dec. 20, 1 945 Axis of impeller -|6 Relative Ar'ea Bladeloading-Ap c. CONCORDIA ET AL 2,484,554

CENTRIFUGAL IMPEL-LE R 2 Sheets-Sheet 2 Fig.9. I

forward Inventors:

Ch ar'les Concordia,

Millar-d' F Dowel Their Attorngs grammatically :sents a modification ofthe impeller in Patented Oct. 11, 1949 2,484,554 CENTRTFUGAL IMPELLERCharles Concordia, Schenectady, N. Y., and Millard F. Dowell,

New York Lynnfield, General Electric Company,

Mass, assignors to a corporation of Application December 20, 1945,Serial No. 636,133 6 Claims. (Cl. 230134) This invention relates toimpellers for turbomachines, particularly to a centrifugal impeller forimparting a pressure rise to a compressible fluid.

The object of the invention is to provide a turbo-machine impeller whichcombines the characteristics of several known types of impellers in amanner to produce improved'efficiency and other performancecharacteristics.

Another object is to provide a turbo-machine impeller, having blades ofa novel shape specifically designed to produce a preselected desiredloading of the blades.

Other objects and advantages will appear from the following descriptionand the appended claims, taken in connection with the accompanyingdrawings, in which Fig. 1 is a perspective view of an impeller havingblades designed in accordance with our invention; Figs. 2, 3 and 4represent top, front and side views of a single blade; Figs. 5, 6, 7 and8 are diagrammatic illustrations of previously known centrifugalimpellers; Fig. 9 is a graphic representation of certain characteristicsof an impeller embodying the invention; and Figs. 10 and 11 representpossible modifications of the basic impeller form shown in Fig. 1.

In order to explain clearly the shape and characteristics of our newimpeller, reference will be made to various types of centrifugalimpellers known to the prior art. Fig. 5 represents diagrammatically themost usual form of centrifugal impeller, having fiat blades eachcomposed of straight-line elements parallel to, and lying in a radialplane through, the axis of rotation. (For simplicity, the impellersrepresented diain Figs. 5, 6, '7 and 8 have been shown with only a fewblades). Fig. 6 repre- Fig. 5, known as a backward sloped impeller. Theblades of this type are characterized by the fact that while they areflat and extend in a substantially radial direction, the plane of theblade does not pass through the axis of rotation but is inclinedbackwardly with respect to the direction of rotation, which direction isindicated by the arrow l8. The type of blade inclination represented inFig. 6 will be referred to herein as "backward slope. Fig. '7 representsanother known type of centrifugal impeller similar to those of Figs. 5and 6, except that the blades are forward sloped." Here each blade isflat, lies in a plane perpendicular to the plane of rotation (that is,parallel to the axis of rotation) but not passing through said axis,being inclined forwardly with respect to the direction of rotationindicated by the arrow l8. Fig. 8 represents an impeller which is wellknown as a "mixed fiow impeller. This type is characterized by the factthat the blades are substantially fiat but the plane of the blade istipped" forwardly in the direction of rotation so that the plane of theblade forms an acute angle with the plane of rotation. The mixed flowimpeller is further characterized by the fact that it has an appreciableaxial depth so that the tipped blades act partly like an axial flowcompressor and partly like a centrifugal compressor. It is because ofthis combination of axial and radial flow that this type has been calleda mixed flow impeller. To recapitulate, the term slope" will be usedherein to denominate the type of inclination shown in Figs. 6 and "I,while the term tip will be used to mean blade inclination as in the caseof the mixed flow impeller of Fig. 8.

We have discovered by analysis, backed by long experience with knownforms of centrifugal impeller, that optimum efliciency and otherperformance characteristics can be obtained from a turbo-machineimpeller if due consideration is given to the blade loading. Bladeloading may be defined as the differential static pressure existingacross an infinitesimal increment of area of the blade at a givenlocation along the fluid flow path through the impeller. The definitionand significance of this factor may be explained by reference to thedrawings, in which Fig. 1 shows an impeller comprising a blade supportbody 8 having a number of equally spaced substantially radially arrangedblades 3. The mean flow path through the fluid passages defined by theblades of the impeller shown in Fig. 1 may be represented by thedot-dash line I, which may be considered to be, the locus of the centersof gravity of cross sections of the flow passage, taken in a directionperpendicular to the mean fiow path at the given section. At the pointI) on the mean flow path i in Fig. 1, the blade loading would berepresented by the difference in the static pressure exerted by theflowing fluid on an infinitesimal increment of blade area represented,to an enlarged scale, by the rectangle 2.

By reference to well-known aerodynamic prinv ciples governing fluid flowin such an impeller, it can be shown that for a blade of a givenpreselected geometric shape, the blade loading can be determined fromthe diiferentialequation in which p is the density in mass units 1',distance of a given axis 01 rotation in particle of fluid from the w,average angular velocity of a particle of fluid I about the axis ofrotation at radius r or at time t du/dt, time rate of change of angularvelocity u, that is, angular acceleration of a fluid particle about theaxis of rotation V7, radial component of the absolute velocity in spaceV of a fluid particle at radius r from the axis of rotation Ap, bladeloading A0, angular-distance between correspondingmean flow paths ofadjacent passages (equal to 21f radians divided by the number ofblades).

Then, noting that density p increases as the static pressure of thefluid increases and that .angular velocity u of a fluid particle at agiven point in the flow passage depends both on the rotational speed ofthe impeller and on the direction of the flow path at the given point,this expression can be solved by various known methods (as for instanceby graphical integration) to obtain the blade loading Ap at the givenpoint.

Then the increment of pressure rise produced by the impeller bladesbetween two points on a given flow path, for instance the points a and bon path I in Fig. 1, can be represented by where in which a. is thevalue of a: at point a dr, differential of the radius r Ap, bladeloading from Equation (1) dp, acute angle which a plane tangent to theblade at a given point makes with a plane defined by the rotation Vz+r,component of the absolute velocity in space of a fluid particle at agiven point, lying in a plane through the axis of rotation and having a2 component in an axial direction and an r-component in a. radialdirection (measured parallel with and perpendicular to the axis ofrotation respectively) dV=+r, diiferential of the velocity component Theabove expressions to suggest one possible will be recognized and methodof analysis and readily understood by those skilled in fluid mechanics.Further details 1 of the method of calculating the design of the bladeare not necessary to the disclosure of our invention.

(in accordance with an increase in distance from given point and theaxis of are given here merely the third term being negligible. Withjection views of one are twenty-two identicalcurved blades. spacedcircumi'erentially' and having portion 4,

the axis of rotation) with a resulting increase in the centrifugal forceexerted on the fluid. Likewise, the second term represents the incrementin pressure (which increment may be positive or negative) due to theangle of slope" of the blades measured to a radial plane through theaxis of revolution. This second term will be positive for a "forwardsloped impeller as in Fig. 7, and negative for a "backward slopedimpeller"'as .in Fig. 6. The third term represents the eiiect of thevariation in flow path area, disregarding all tangential components offluidmotion and considering only the difiusing eliect of the flowpassage, that is, the conversion of kinetic energy third term alsobecomes important. It maybe noted in passing that for a pure axial flowcompressor (not shown in the drawings) the above expression for theblade loading is substantially the second term plus the third term, withthe first term. having only an incidental effect.

A consideration layer characteristics of centrifugal impellers led us tothe conclusion that best efliciency could .be-

obtained if the blade loading, as defined above,

could be made substantially zero at the inlet tothe impeller and causedto increase smoothly but rapidly to a maximum value which is maintainedover-a major portion of the flow path and then caused to drop smoothlybut rapidly to substantially zero at the exit of the im peller flowpassages. We have found that a composite impeller combining in aparticular way the characteristics of the'mixed flow impeller of Fig. 8,the forward sloped impeller of Fig. 7, and the backward sloped impellerof Fig. 6, produces such a blade loading and gives superior performance.The manner in which these three known types of impeller blade shapes arecombined will be readily apparent from the perspective view of acomplete impeller shown in Fig. 1, taken in connection with theorthographic problade shown in Figs. 2, 3, and .4.-

' Fig. 1 illustrates a complete impeller, consisting diameter at oneaxial end and a minor diameter at the other axial end Ythe axial lengthof the body being approximatelyof the same magnitude as the majorradius. This-body has a concave outer surface (appr ximately ahyperboloidal surface) on which the blades aresecured and is providedwith an axial bore 9 for mounting the impeller on a shaft. Supported onthe body 8 3, equally an inlet an intermediate portion 5, and a tip ordischarge portion 6. The extent of each of these threev distinct bladeportions is clearly indicated in Figs. 1-4. Inlet part 4 extends fromthe inlet blade impeller as in Fig. 5,

(but usually negative), while the of the fundamental. boundarya solidbody of revolution 8 having a major.

edge I to a point approximately one-third the length of the mean flowpath I from the inlet edge I. This corresponds to the region of rapidchange of curvature indicated by the shade lines near the middle of theblade in Fig. 3. Likewise the intermediate portion merges into the tipportion 6 at the location indicated by the shade lines adjacent the:blade tip in Fig. 3.

From a consideration of Figs. 1-4, it will be readily apparent that theinlet portions I are in the form of mixed flow impeller blades, similarto those represented in Fig. 8; the intermediate blade portions 5 are inthe form of forwardly sloped impeller blades, as shown in Fig. '7; whilethe tip portions 6 are in the shape of backwardly sloped impellerblades, as in Fig. 6. These blade portions must of course be modifiedslightly at the place where they join an adjacent portion so that themean flow path through the fluid passages formed between adjacent bladeswill be a smooth curve.

The inlet portion 4 is designed in accordance with the well-knownprinciples governing the design of the mixed flow impeller so that fluidapproaching the inlet edge 'I of the blades, in a substantially axialdirection, is smoothly accelerated by the blade in a tangential andaxial direction, and perhaps also given some very slight acceleration ina radial direction. The intermediate blade portion 5 defines roughly 50per cent of the length of the mean flow path and furnishes the majorportion of the work energy imparted by the impeller to the fluid. Thecomparatively short tip portion 6 serves to gradually relieve theloading, to a value approaching zero at the exit edge It.

For convenience, our new blade shape may be referred to as an S-blade",because of its characteristic reverse curvature.

The upper portion of Fig. 9 represents graphically the blade loading ofour S-blade impeller as a function of the radial distance from the axisof rotation. While Figs. 2, 3 and 4 are to substantially the same scaleas Fig. 1, the abscissa of Fig. 9 is to approximately double the scaleof Figs. 1-4. The heavy curve shown represents the loading along themean flow path I. It will be understood by those skilled in the art thatsimilar loading curves can be calculated for flow paths occupying otherpositions in the fluid passage. for instance one lying along the forwardportion of the fluid passage (relative to the direction of rotation)which maybe represented by the line I I in Fig. 1. Likewise the bladeloading can be calculated for a flow path lying along the r arwardportion of the fluid passage, as for instance that represented by theline I! in Fig.

4 1. Similarly, load curves can be calcualted for flow paths lyingadjacent the curved outer surface of the body 8, or adjacent the curvedinner surface of the shroud which forms the fourth wall of the fluidflow passages.

No shroud has been illustrated for the impeller of Fig. l; but it willbe readily understood by those skilled in the art that the impellershown is what is known as an open impeller which, when assembled in itscasing. is closely surrounded. by a stationary curved wall forming aclose clearance with the free edges l3 of the blades. An open impellerassembled with its casing and other component parts is shown in U. S.Patent 2,377,740, filed March 31, 1944 in the name of J. S. Alford. Itwill be obvious that our invention may be incorporated also in ashrouded or "closed" type impeller, in which the fourth wall 'iiuid flowpath, deflned by the inlet edges I of the blades, occurs at a radiusroughly 25 per cent of the tip radius. The termination of the mixed flowor inlet portion 4 of the blade corresponds roughly to point It in Fig.9, at a radius of about 40 per cent of the tip radius. The loadingstarts at zero, or a very low initial value, at inlet edge I; and by thetime the fluid reaches point ll, the blade loading has increased to verynearly its maximum value, which is maintained at a high, substantiallyuniform, value throughout the intermediate. blade portion 5. Theintermediate portion terminates approximately at point l5; and from thatpoint to the blade tip II), the loading decreases smoothly but rapidlyto substantially zero, under the influence of the backward sloped tipportion 6.

For purposes of comparison, curves are shown in Fig. 9 representingtypical blade loadings for the known types of impeller represented inFigs. 5-8. The forward sloped impeller gives the highest blade loading;the backward sloped impeller gives much lower loadings; while the plainradial blade produces a loading curve lying between those for theforward sloped and backward sloped impellers respectively. The mixedflow impeller has a distinctive blade loading curve quite different fromthe other known types represented. .In producing our new blade shape wehave utilized the characteristics of the forward sloped blade to keepthe blade loading uniformly high over the intermediate portion of theblade, which does the major portion of the work. The characteristics ofthe mixed flow impeller are employed to increase the loading from a lowinitial value to a value in the neighborhood of that produced by theintermediate portion. By merging the forward sloped intermediate portionsmoothly with the backward sloped tip portion, the loading curve iscaused to drop smoothly to a very low value, or to zero.

We have discovered by analysis and experiment that optimum results areobtained when the mixed flow portion 4 of our S-blade defines from 20 to40 per cent of the mean fluid flow path, while the backward sloped tipportion 6 defines from 10 to 30 per cent of the mean flow path. It hasalso been found that best results are obtainedwhen the length componentof the mean flow path in a direction parallel to the axis of rotation isfrom 40 to 60 per cent of the outside diameter of the impeller, that is,when the axial depth of the impeller is roughly equal to the tip radiusof the blades. For convenience, this axial length component of the fluidflow passages may be referred to as the Z-dimension. This is the same ewhich appears in Equation (2) above.

It is well known to those skilled in this art that energy losses due toturbulence are likely to occur when the shape of the walls defining thefluid flow passage is such as to cause a change in direction of thefluid. This tendency becomes more pronounced as the rate of curvatureincreases, by reason of separation of the boundary layer from thepassage wall. It has been found that such turbulence losses can bereduced by causing the fluid flow passage to converge slighty concurrentwith the change in direct on. In accordance with this principle, theflow passages defined by our S-shaped blades decrease in crosssectionalarea as represented by the curve shown in the lower part of Fig. 9. Itwill be seen that the cross-sectional area of the passage progressivelydecreases throughout the mixed flow portion 4, in which part of the flowpath the maximum change in direction of the fluid flow occurs. Thecross-sectional area of the intermediate and tip portions of the flowpassage may remain substantially constant, as represented by the curvein Fig. 9, or may continue to decrease slightly, or may in someimpellers increase again. We have found that best results are obtainedif the passage area decreases to a value between 70 and90 per cent ofthe area at the inlet to the impeller by the time the heavily loadedintermediate portion 5 of the blade is reached. The relative area curveshown in Fig. 9 reaches a minimum of approximately 80 per cent of theinlet area at a location slightly beyond the point M on the bladeloading curve in Fig. 9.

The relative cross-sectional area of the fluid flow passages through ourimpeller may be varied either l) by reducing the height of the passage(measured in a direction normal to the surface of i the blade supportbody 8), or (2) by thickening the blades so'as to reduce the width ofthe passage (measured from one blade to the next adjacent blade in adirection perpendicular to the mean flow path). Either or both of thesemethods, may be used in making this relative cross-sectional area of theflow path vary in the desired manner. Consideration of Fig. 1 willreveal that the blades are thickest in the region where the mixed flowblade portion 4 merges into the forward sloped portion 5. Thisthickening of the blades is an important factor intended to reduce thecross-sectional area in accordance with the relative area curve in Fig.9.

Figs. 1-4 illustrate our basic blade shape with the blades arrangedrelative to each other and to the support body 8 so that the arbitrary"center-line" i1 shown in Fig. 2 passes through the axis it of theimpeller. Figs. 10 and 11 show possible modifications, in which thecenter-line l'l is "sloped relative to the direction of rotation(indicated by' the arrow It). In Fig. 1 blade 3 is "backwardly' slopedby the amount of angle l9; that is, the center-line l'lforms the angleis witha' radial plane. In Fig. 11 the arbitrary center-line l l'isforwardly sloped, making the angle 20 with a radial plane.

Sloping the blade backwardly'as'T in Fig. 10 has the effect ofdecreasing the average blade loading; while sloping the blade'iorwardlyas in Fig. 11 has the efiectof increasing the blade loading. If ourS-shaped blade is sloped backwardly far enough, in the manner of Fig. 10but to a greater degree, the intermediate portion of the blade maybecome substantially radial. Likewise, if the blade is sloped forwardlya suficient amount, in the manner of Fig. 11, the discharge portion 6 ofthe blade may closely approach the radial direction. Many suchmodifications have been studied by us, and all have their place in thedesign of various impellers of special characteristics.

While our improved turbo-machine blade shape has been specificallydescribed as applied to a compressor impeller, it will be, apparent tothose skilled in the art that the invention is also applicable toturbine rotors, in -which case nozzles would be arranged todirect-motive fluid radially, inward to the blade tip portions- 6. Thusthe fluid would flow through the impeller in the reverse direction, ascompared with the flow in a compressor impeller; and the pressuregradients, blade loadings, etc. would be similar qualitatively to thoseobtaining in a compressor impeller incorporating our invention. a

What we claim as new and desire to secure by Letters Patent of theUnited States, is:

1. In a compressor impeller arranged to receive fluid with asubstantially axial velocity adjacent the axis of rotation and todischarge it at a ma terially greater radius from the axis of rotation,walls defining a passage having a mean flow path consisting of a firstportion forming 20 to 40 per cent of the length of the path nearest theaxis of rotation and tipped forwardly in the direction of rotation toform an acute angle with the plane of rotation, a second portion forming10 to 30 per cent of the length of the flow path most remote from theaxis of rotation and sloped backwardly relative to the axis, and aportion intermediate the first and second portions forming the remainderof the flow path and sloped forwardly relative to the axis, said passagehaving a cross-sectional area which progressively decreases from theinlet of the first portion to a value of from 70 to per cent of theinlet area at the juncture of the first portion with the inter mediateportion.

2. In a centrifugal impeller arranged to receive fluid with asubstantially axial velocity adjacent the axis of rotation and todischarge it at a materially greater radius from the axis of rotation,walls defining a fluid flow passage, said walls including a plurality ofcurved blades circumferentially spaced from each other, each bladehaving an inlet portion in the form of a mixed flow impeller bladedefining 20 to 40 per cent of the length of the fiow path from the inletthereto, a

discharge portion in the form of a backward sloped impeller bladedefining 10 to 30 percent of the length of the flow path adjacent theexit thereof, and an intermediate portion in the form of a forwardsloped impeller blade defining the remainder of the flow path connectingthe inlet portion and the discharge portion, the width and thickness ofsaid blades being dimensioned so that the cross-sectional area of thefluid passagedefined by adjacent blades progressively decreases from theinlet of the mixed flow portion to a value of from 70 to 90 per cent ofthe inlet area at the juncture of the mixed flow and forward slopedportions.

3. A turbo-machine impeller comprising a blade support body having asubstantially hyperboloidal outer surface and a plurality ofcircumferentially spaced radially extending blades secured to saidsurface, each adjacent pair of blades defining a fiow path having afirst radially inner terminal portion constituting 20 to 40 per cent ofthe total length of the flow path and having a mean fiow pathsubstantially straight with a small radial component and substantialtangential and axial components, a second radially outer terminalportion forming 10 to 30 per cent of the total length of the flow pathand having a mean flow path substantially straight with a negligibleaxial component and a substantial radial component, and an intermediateportion connecting the first and second portions and having a mean flowpath substantially straight with substantial axial and radialcomponents, the intermediate portion being modified at its respectiveends to cooperate with the first and second portions so as to form asmooth reversely curved flow path having a, blade loading which is.maintained at a substantially uniform high value throughout theintermediate portion of the flow path and falls off smoothly throughoutthe first and second terminal portions to minimum values approachingzero at the impeller inlet and exit respectively.

4. A turbo-machine impeller comprising a blade support body having asubstantially hyperboloidal outer surface and a plurality ofcircumferentially spaced radially extending blades secured to saidsurface, each adjacent pair of blades defining a flow path having afirst radially inner terminal portion constituting 20 to 40 per cent ofthe total length of the flow path and having a mean flow pathsubstantially straight with a small radial component and substantialtangential and axial components, a second radially outer terminalportion forming to 30 per cent of the total length of the flow path andhaving a mean flow path substantially straight with a negligible axialcomponent and a substantial radial component, and an intermediateportion connecting the first and second portions and having a mean flowpath substantially straight with substantial axial and radialcomponents, the blades being dimensioned so that the cross section areaof the fiow path defined therebetween decreases progressively throughoutthe first portion to a value at the juncture with the intermediateportion of from 70 to 90 per cent of the area at the other end of thefirst portion, the intermediate portion being modified at its respectiveends to cooperate with the first and second portions so as to form asmooth reversely curved flow path having a blade loading which ismaintained at a substantially uniform high value throughout theintermediate portion of the flow path and falls off smoothly throughoutthe first and second terminal portions to minimum values approachingzero at the impeller inlet and exit respectively.

5. A turbo-machine impeller comprising a blade support body having asubstantially hyperboloidal outer surface and a plurality ofcircumferentially spaced radially extending blades secured to saidsurface, each adjacent pair of blades defining a flow path having afirst radially inner terminal portion constituting 20 to 40 per cent ofthe total length of the flow path and having a mean fiow pathsubstantially straight with a small radial component and substantialtangential and axial 60 components, a second radially outer terminalportion forming 10 to 30 per cent of the total length I of the flow pathand having a mean fiow path substantially straight with a negligibleaxial component and a substantial radial component, and an intermediateportion connecting the first and second portions and having a mean flowpath substantially straight with substantial axial and radialcomponents, the intermediate portion being modified at its respectiveends to cooperate with the first and second portions so as to form asmooth reversely curved flow path having a blade loading which ismaintained at a substantially uniform high value throughoutthe,intermediate portion of the flow path and falls off smoothlythroughout the first and second terminal portions to minimum valuesapproaching zero at the impeller inlet and exit respectively, the axiallength of the complete flow path being from 40 to per cent of the outertip diameter of the path.

6. In a centrifugal impeller adapted to receive fluid with asubstantially axial velocity adjacent the axis of rotation and todischarge it at a materially greater radius from the axis of rotation,walls defining a fluid flow passage having an inlet portion the meanflow path of which has a substantial tangential component in thedirection of rotation, said walls including a blade support body havinga substantially hyperboloidal outer surface and a plurality ofcircumferentially spaced radially extending blades secured to said body,the width and thickness of said blades being so dimensioned that thecross-sectional area of the fiow passage decreases progressivelythroughout the inlet portion to a value at the end of said portion ofthe order of per cent of the inlet area, whereby turbulence and boundarylayer separa tion are reduced.

CHARLES CONCORDIA. MILLARD F. DOWELL.

REFERENCES CITED The following references are of record in the file ofthis patent;

UNITED STATES PATENTS Number Name Date 673,244 Davidson Apr. 30, 19011,314,049 Criqui Aug. 26, 1919 1,959,703 Birmann May 22, 1984 2,399,852Campbell et al. May 7, 1946

